1. (Field of the Invention)
The present invention relates to a combined externally pressurized gas and magnetic bearing assembly in which an externally pressurized gas bearing and a magnetic bearing are combined together, and also to a spindle device employing the combined externally pressurized gas and magnetic bearing assembly. More particularly, the present invention relates to the combined externally pressurized gas and magnetic bearing assembly and the spindle device both suited for use in, for example, a high speed milling machine.
2. (Description of the Prior Art)
The magnetic bearing is known to have a feature in that because it has a relatively large bearing gap a loss of torque during rotation thereof is extremely small and it can exhibit a high static rigidity by an integral control.
FIG. 28 illustrates a longitudinal sectional view showing a prior art spindle device utilizing a magnetic bearing, which is used in a high-speed aluminum milling machine. This prior art spindle device includes a touch-down bearing 251, a tool 252, a displacement sensor 253, a radial magnetic bearing 254, an axial magnetic bearing 255, a drive motor 256, a radial magnetic bearing 257, a displacement sensor 258, and a main shaft 259. The spindle device utilizing the magnetic bearings has a performance of a maximum rotation of 40,000 rpm, and a maximum output of 15 kW, a maximum machining capacity of 1,250 cm3/min, thus exhibiting an excellent performance for an aluminum milling work.
However, the spindle device utilizing the magnetic bearing is susceptible to influence brought about by a natural frequency of bending of the main shaft during the milling operation and, for this reason, requires the use of an extremely complicated control system. Accordingly, the known spindle device of the type discussed above is not suited as a spindle device for a versatile machine tool that is required to accommodate various processing conditions.
On the other hand, a non-contact bearing currently available other than the magnetic bearing includes an externally pressurized gas bearing. Although the externally pressurized gas bearing is known to have a high rotational accuracy and an excellent dynamic stability, the externally pressurized gas bearing has been little used in the versatile machine tool because of the compressivity and, hence, a low static rigidity and a low load bearing capacity.
In view of the foregoing, attempts have recently been made to use, as a spindle device for high speed machining purpose, a spindle device utilizing a composite bearing in which the externally pressurized gas bearing and the magnetic bearing are combined, such as shown in FIG. 29 in a longitudinal sectional representation. Referring to FIG. 20, reference numeral 263 represents a displacement sensor; reference numeral 264 represents a radial magnetic bearing; reference numeral 265 represents an axial magnetic bearing; reference numeral 266 represent a drive motor; reference numeral 267 represents a radial magnetic bearing; reference numeral 268 represents a displacement sensor; reference numeral 270 represents a displacement sensor; reference numeral 27 represents a main shaft; and reference numerals 272 and 273 represent respective externally pressurized gas bearings.
However, the spindle device utilizing the composite bearing shown in FIG. 29 has a problem in that since the magnetic bearings 264 and 267 and the externally pressurized gas bearings 272 and 273 are disposed one after another in an axial direction of the main shaft 271, not only is a relatively long main shaft 271 required, but the natural frequency of bending tends to be lowered. Also, since the spindle device shown in FIG. 29 makes use of a control system of a structure similar to that required in the spindle device utilizing solely the magnetic bearings, an additional problem has arisen that the dynamic stability of the externally pressurized gas bearing is impaired and it tends to function as a source of external disturbances.
Accordingly, the art has not yet been developed to the extent as to fulfill the objective of utilizing advantages of the externally pressurized gas bearing and also those of the magnetic bearing while counterbalancing demerits of those respective bearings.
Accordingly, the present invention has been devised with a view substantially eliminating the above discussed problems and is intended to provide a novel combined externally pressurized gas and magnetic bearing assembly which has an excellent dynamic rigidity, such as exhibited by the externally pressurized gas bearing, and also an excellent static rigidity, such as exhibited by the magnetic bearing, and which contributes to reduction in length of the main shaft, and an -improved spindle device utilizing the novel combined externally pressurized gas and magnetic bearing assembly.
In order to accomplish the foregoing and other objects of the present invention, there is provided a novel combined externally pressurized gas and magnetic bearing assembly comprising an externally pressurized gas bearing and a magnetic bearing combined together in a predetermined relationship. Because of this, the bearing assembly can be obtained which can exhibit both the dynamic rigidity possessed by the externally pressurized gas bearing and the static rigidity possessed by the magnetic bearing.
While numerous combined externally pressurized gas and magnetic bearing assemblies of different constructions within the spirit of the present invention are disclosed, the combined externally pressurized gas and magnetic bearing assembly according to a first aspect of the invention is so designed as to combine the externally pressurized gas bearing and the magnetic bearing together by providing a magnetic bearing including an electromagnet having a core and a fluid restrictor defined in the core for supply of a gaseous medium under pressure into a bearing gap of the externally pressurized gas bearing.
The combined externally pressurized gas and magnetic bearing assembly according to a second aspect of the invention is of a design in which the magnetic bearing and the externally pressurized gas bearing are combined together with the magnetic bearing so as to have commonly shared component parts. It is to be noted that the combined externally pressurized gas and magnetic bearing assembly according to any one of the first to third aspects of the invention may be used as either a radial bearing or an axial bearing.
The combined externally pressurized gas and magnetic bearing assembly according to a third aspect of the invention is of a design in which a fluid restrictor is formed by a self-forming orifice or a normal orifice in a bearing stator of the magnetic bearing. The self-forming orifice performs a function as an orifice at a virtual cylindrical surface defined by and between a gas conduit and a bearing gap. The magnetic bearing includes a displacement measuring means for measuring a displacement of a rotor and is operable to support the rotor at a predetermined position by generating an electromagnetic force according to the displacement measured by the displacement measuring means. The displacement measuring means may be of a type capable of directly measuring the displacement of the rotor or of a type capable of indirectly measuring the displacement of the rotor, that is, of a type capable of measuring a different object to be measured that can eventually be converted into the displacement of the rotor.
The combined externally pressurized gas and magnetic bearing assembly according to a fourth aspect of the invention is a radial bearing and is of a design in which the externally pressurized gas bearing is accommodated within a width of the magnetic bearing in an axial direction.
The combined externally pressurized gas and magnetic bearing assembly according to a fifth aspect of the invention is a radial bearing and is of a design in which the magnetic bearing is accommodated within a width of the externally pressurized gas bearing in an axial direction.
The combined externally pressurized gas and magnetic bearing assembly according to a sixth aspect of the invention is a radial bearing and is of a design in which a bearing gap of the externally pressurized gas bearing a gap between the rotor and a core of the magnetic bearing are defined at the same position in an axial direction.
Where the magnetic bearing and the externally pressurized gas bearing are integrated together such as set forth in any one of the first to sixth aspects of the invention, the bearing assembly as a whole can be assembled compact as compared with the bearing assembly in which an externally pressurized gas bearing and a magnetic bearing are merely laid one after another in the axial direction.
Where the combined externally pressurized gas and magnetic bearing assembly is used as a radial bearing, the main shaft which serves as the rotor does not require an extra length solely for the purpose of accommodating the support by an externally pressurized gas and the support by a magnetic force separately and, therefore, the bearing assembly can have a reduced length in the axial direction with the main shaft of a reduced length. Accordingly, the natural frequency of bending can be increased to accomplish a high speed rotation. Also, the center point of support by the magnetic bearing and the center point of support of the externally pressurized gas bearing can be substantially matched with each other in the axial direction and, therefore, a control of both of those bearings can be achieved easily.
Where the combined externally pressurized gas and magnetic bearing assembly is used as an axial bearing, the bearing assembly as a whole can be assembled compact as compared with the bearing assembly in which an externally pressurized gas bearing and a magnetic bearing are merely laid one after another in the radial direction and, therefore, the diameter of each of opposite surfaces of the rotor confronting the bearings can be reduced.
Also, where the self-forming orifice is used for a fluid supply system of the externally pressurized gas bearing, the stability against the pneumatic hammer can be increased together with increase in bearing stability at the high frequency region, that is, in dynamic rigidity.
In any event, the self-forming orifice of the externally pressurized gas bearing is preferably defined in an electromagnetic force generating surface of the core of the electromagnet in the magnetic bearing and a gap between the electromagnetic force generating surface and the rotor is preferably of a size equal to or smaller than 0.1 mm.
The use of the fine self-forming orifice in the core of the electromagnet is effective to maximize the utilization of the cross-section of the core of the electromagnet which has hitherto been utilized only to provide a magnetic circuit for magnetic fluxes of the magnetic bearing. A space between neighboring yokes of the core can be utilized as a space for accommodating coils of the electromagnet and, therefore, although where the orifice is to be formed in that space, interference with the coils of the electromagnet must be taken into consideration, the formation of the orifice in the core eliminates such a problem associated with the interference with the coils.
In the event that the bearing assembly of the foregoing construction is used as a radial bearing, three or more electromagnets forming the magnetic bearing may be arranged in a circumferential direction with a pair of magnetic poles of each of the electromagnets oriented towards a rotary shaft in such a way that polarities of the magnetic poles of each electromagnet arranged on the same circumference may be matched with each other. The core of each electromagnet may be provided in a shape generally similar to, for example, a horseshoe, in a longitudinal cross-section.
By constructing the core of each electromagnet in the manner described above, the loss of hysteresis and the loss of eddy current which would be brought about at the rotor of the magnetic bearing incident to rotation of the rotor can be reduced advantageously. Because of those losses being lessened, emission of heat from the rotor can be suppressed and, therefore, reduction in size of the bearing gap which would be brought about by a thermal expansion of the rotor can be minimized to allow the externally pressurized gas bearing to exhibit a stabilized performance.
In the structure in which the magnetic poles are arranged in a direction parallel to a rotary shaft, ones of the magnetic poles of the cores of all of the electromagnets are mutually linked with each other.
By so constructing the core, not only can the number of steps of a process of making the electromagnet be reduced, but also the core loss at the rotor of the magnetic bearing which would occur incident to rotation of the rotor can further be reduced to permit the bearing assembly to accommodate a high speed rotation.
Where the combined externally pressurized gas and magnetic bearing assembly of the structure as set forth in any one of the first to third aspects of the invention is used as an axial bearing, the electromagnet and the fluid restrictor of the externally pressurized gas bearing may be arranged only on one side of the rotor with respect to the axial direction so that the rotor can be supported by an electromagnetic force of the magnetic bearing and a bearing force of the externally pressurized gas bearing that is counterbalanced with the electromagnetic force. The use of the electromagnet and the fluid restrictor only on one side of the rotor renders the bearing assembly to be assembled further compact in size.
The present invention also provides a spindle device utilizing the combined externally pressurized gas and magnetic bearing assembly. The bearing assembly used in the spindle device may be of the structure according to any one of the foregoing aspects of the invention for rotatably supporting a main shaft having the rotor formed therein.
According to this spindle device, not only can a relatively high static rigidity and a relatively high dynamic stability be obtained, but the main shaft can have a reduced length with increase in natural frequency of bending of the main shaft and, therefore, the spindle device can be used as a versatile high-speed machining spindle device.
In this spindle device of the structure described above, the combined externally pressurized gas and magnetic bearing assembly may be mounted on a load-side end portion of the main shaft. This is effective to a bearing force to effectively act against a radial load transmitted to that end portion of the main shaft. In particular, with respect to the static load, a high rigidity can be obtained by allowing the control of the magnetic bearing to be performed on the basis of an integral action or a proportional integral action.
The displacement measuring means for measuring the displacement of the rotor, which is employed in the combined externally pressurized gas and magnetic bearing assembly of the structure may be employed in the form of a pressure sensor for measuring a pressure developed at a bearing surface of the externally pressurized gas bearing so that the magnetic force developed by the magnetic bearing can be controlled by determining the displacement of the rotor by the utilization of a value measured by the pressure sensor.
By directly measuring the pressure at the bearing surface of the externally pressurized gas bearing and converting the measured value into the displacement of the rotor which is subsequently used to control the bearing, any possible erroneous operation of the sensor which would occur as a result of variation in magnetic characteristic at a rotor sensor target surface as is often observed with a sensor of any other system can be eliminated to accomplish a highly accurate sensing. By way of example, not only can detection of the displacement be achieved with high accuracy, for example, with a resolving power of the sub-micron order, but the sensor can be made compact enough to accommodate in the bearing assembly thereby rendering the bearing assembly compact in size. Also, by converting the pressure of the externally pressurized gas bearing into the displacement of the rotor, the center of bearing of the externally pressurized gas bearing can be set at the center of bearing of the magnetic bearing and, therefore, any possible interference therebetween can easily be avoided.
Where the pressure sensor is employed in the form of a semiconductor pressure sensor, a pressure measuring unit can be fabricated compact in size.
On the other hand, where the combined externally pressurized gas and magnetic bearing assembly of the structure is used as a radial bearing, the pressure sensor may be used to measure a difference of a static pressure at externally pressurized gas bearing surfaces opposed to each other in a direction diametrically of the rotor so that the displacement of the rotor can be determined by utilizing the values measured by the pressure sensor.
By so arranging the pressure sensor, the displacement of the rotor can be determined with high accuracy with a minimized number of pressure sensors used.
Where this combined externally pressurized gas and magnetic bearing assembly is used as an axial bearing, the pressure sensor may be used to measure respective pressures at three or more locations of the externally pressurized gas bearing surface, which are positioned on the same circumference of the externally pressurized gas bearing surface, so that the measured pressure values can be utilized to determine the displacement of the rotor in an axial direction. By setting the locations at which the pressure is to be measured, the displacement of the rotor in the axial direction can be measured accurately with a minimized number of the pressure sensors and, and hence, at a reduced cost.
Instead of the three locations at which the pressure is measured, respective pressures at two locations, which are positioned diametrically opposite to each other, may be measured by the pressure sensor so that the measured pressure values can be utilized to determine the displacement of the rotor in the axial direction. By so doing, the displacement of the rotor can be determined with high accuracy with a minimized number of pressure sensors used.
Where the combined externally pressurized gas and magnetic bearing assembly of the structure described above is used as an axial bearing, the electromagnets of the magnetic bearing and the fluid restrictor of the externally pressurized gas bearing may be positioned on respective axially opposite sides of the rotor. In such case, the pressure at an arbitrary location of one of externally pressurized gas bearing surfaces opposite to each other with respect to the rotor and the pressure at a location of the other of the externally pressurized gas bearing surfaces which is symmetrical to said arbitrary location about a point on the rotor may be measured by the pressure sensor to provide two measured values which are utilized to determine the displacement of the rotor in an axial direction.
By so setting the locations at which the pressure is measured, the displacement of the rotor in the axial direction can be measured accurately with a further minimized number of the pressure sensors and, and hence, at a further reduced cost.
In the combined externally pressurized gas and magnetic bearing assembly of any of the foregoing structures, the pressure sensor may positioned without being spaced from a pressure measuring point on the externally pressurized gas bearing surface. In other words, the pressure sensor may be fixed at or adjacent a location at which the pressure at the bearing surface of the externally pressurized gas bearing assembly is to be measured. By so positioning the pressure sensor, the pressure at the desired pressure measuring surface can be measured directly.
Also, in the combined externally pressurized gas and magnetic bearing assembly of any of the foregoing structures, instead of the pressure sensor being positioned at the measuring surface, the pressure sensor may be positioned spaced from the externally pressurized gas bearing surface, in which the pressure sensor and the externally pressurized gas bearing surface are to be communicated with each other by means of a micro-passage, a pipe or a combination of the micro-passage and the pipe. In this arrangement, the micro-passage, for example, is defined at the bearing surface of the externally pressurized gas bearing and is in the form of a hole of a diameter equal to or smaller than 1 mm. The pipe may be fluid-connected with the micro-passage and has an inner diameter equal to or smaller than 1 mm. The pressure sensor is fitted to one end of the pipe.
Where bearing surface of the externally pressurized gas bearing or the combined externally pressurized gas and magnetic bearing assembly itself is so small that the pressure sensor cannot be accommodated, or where an empty space is available within the spindle device utilizing the bearing assembly, the use of the micro-passage and/or the pipe is effective. Also, by choosing the diameter of the micro-passage defined at the bearing surface of the externally pressurized gas bearing for the pressure measurement to be equal to or smaller than 1 mm, any possible adverse influence on the externally pressurized gas bearing can be minimized and, by choosing the diameter of the pipe coupled with the micro-passage to be equal to or smaller than 1 mm, the pressure measurement is possible with no need to reduce the frequency characteristic.
In the combined externally pressurized gas and magnetic bearing assembly of any of the foregoing structure, the core of the electromagnet of the magnetic bearing may utilize a solid material. The use of the solid material for the core makes it easy to form the fluid restrictor such as the self-forming orifice and, therefore, the highly precise combined externally pressurized gas and magnetic bearing assembly can be assembled.
Instead of the core of the electromagnet being made of the solid material in its entirety, the core may have a portion thereof made of the solid material while the remaining portion of said core is made of a laminated silicon steel plate. The use of the laminated silicon steel plate makes it possible to reduce the core loss occurring at the core of the electromagnet while the use of the solid material facilitates formation of the fluid restrictor such as the self-forming orifice in the externally pressurized gas bearing.
Again, in the combined externally pressurized gas and magnetic bearing assembly of any of the foregoing structures, the rotor may utilize a laminated silicon steel plate and having a coating layer of a ceramic material formed on the eliminated silicon steel plate to a thickness equal to or smaller than 1 mm.
If the rotor is made of the laminated silicone steel plate, the core loss at the high speed rotation can be lessened to suppress emission of heat from the rotor during the high speed rotation thereof. In addition, the coating of the ceramic material on an outer peripheral surface of the rotor minimizes any possible damage to the rotor which would occur when the rotor is brought into contact with the bearing surface. Formation of the coating layer made of the ceramic material is effective to eliminate the occurrence of the core loss resulting from the magnetic fluxes emanating from the electromagnet of the magnetic bearing and hence, effective for the high speed rotation. Outer and inner peripheral surfaces of the coating layer serve a rotor surface of the externally pressurized gas bearing and a rotor surface of the magnetic bearing, respectively, and by adjusting the thickness of the coating layer, an optimum bearing gap of the externally pressurized gas bearing and an optimum bearing gap of the magnetic bearing can be fixed.
Alternatively, the rotor may utilize a soft magnetic solid material having a low thermal expansion coefficient, in which case the coating layer of the ceramic material is formed on the solid material to a thickness equal to or smaller than 1 mm. Where the rotor is provided on the main shaft, the main shaft is preferably made of the same solid material as the rotor. An example of the solid material that can be employed in the present invention includes invar.
The use of the solid material for the rotor is effective to increase the natural frequency of bending of the rotor to such an extent as to permit the rotor to be driven at a further high speed. In addition, even when the rotor is heated, a change of the bearing gap can be kept minimum because of the low thermal expansion coefficient exhibited by the solid material and, therefore, the stabilized performance of the externally pressurized gas bearing can be secured. Yet, since the amount of thermal expansion in the axial direction is minimum, the machining accuracy can be effectively increased if it is used in the main shaft of a machine tool.
In the combined externally pressurized gas and magnetic bearing assembly of any of the foregoing structures, a magnetic bearing control means may be employed for controlling the magnetic bearing on a feed-back scheme according to the value measured by the displacement measuring means, which magnetic bearing control means is operable to perform the control on an integral action or a proportional integral action and to effect no control to a high frequency higher than a predetermined value.
By performing the control on an integral action or a proportional integral action and effecting no control to a high frequency higher than a predetermined value, the magnetic bearing can be limited to exert the bearing force at the low frequency region without the dynamic stability, which is an advantage of the externally pressurized gas bearing, being adversely affected and, therefore, the static rigidity peculiar thereto can be increased. In other words, the externally pressurized gas bearing and the magnetic bearing can exhibit a dynamic rigidity (a high frequency region) and a static rigidity (a low frequency region), respectively, and therefore, the respective advantages of the magnetic bearing and the externally pressurized gas bearing can be retained to avoid any possible interference therebetween.
Where such magnetic bearing control means is employed, the use may be made of a linearizing circuit in an amplifier unit for processing an output from the displacement measuring means, so that a characteristic relationship between a control voltage and an electromagnetic force can be linearized without supplying a bias current to coils of an electromagnet of the magnetic bearing. The linearization of the characteristic relationship between the control voltage and the electromagnetic force can be accomplished by providing the amplifier unit with a current squaring feedback circuit.
By accomplishing the linearization with no need to supply the bias current, no negative rigidity peculiar to the magnetic bearing will be generated.
Also, a means for processing the value measured by the displacement measuring means may have an insensitive zone, so that the magnetic bearing control means will not control the magnetic bearing in the event that the displacement measured by the displacement measuring means is within a predetermined range. Means for providing the insensitive zone may be provided either in a front stage of the magnetic bearing control means separately from the magnetic bearing control means, or in the magnetic bearing control means.
In the event that the center of the magnetic bearing and the center of the externally pressurized gas bearing deviate slightly from each other because of a faulty adjustment of the displacement sensor, or in the event that a drift occurs in the displacement sensor in dependence on the temperature or the like, even though the rotor is kept in alignment with the center of the externally pressurized gas bearing, an electric current will flow through the coils of the magnetic bearing, thereby constituting an external disturbance to the externally pressurized gas bearing. The insensitive zone is utilized to eliminate this problem to thereby eliminate any possible interference from the magnetic bearing to the externally pressurized gas bearing, wherefore the stabilized combined externally pressurized gas and magnetic bearing assembly can be assembled.
In the case of the spindle device utilizing the combined externally pressurized gas and magnetic bearing assembly for rotatably supporting the main shaft having the rotor formed thereon, a control means may be used by which after the main shaft has been floated on a non-contact basis upon activation of the externally pressurized gas bearing during a start of the externally pressurized gas bearing, a direct current component of an output from the displacement measuring means being zeroed, followed by activation of the magnetic bearing. This control means may be implemented either by the magnetic bearing control means or by a means separate therefrom.
By driving this way, the temperature dependent drift of the sensor which serves as the displacement measuring means in the spindle device where the temperature tends to increase can be compensated for to avoid any possible erroneous operation of the magnetic bearing which exerts the electromagnetic force based on an output from the sensor.
Also, in the spindle device utilizing the combined externally pressurized gas and magnetic bearing assembly for rotatably supporting the main shaft having the rotor formed thereon, a control means may be used by which after the main shaft has been driven and floated on a non-contact basis upon activation of the externally pressurized gas bearing during a start of the externally pressurized gas bearing, a direct current component of an output from the displacement measuring means being zeroed when the main shaft attains a predetermined number of revolutions or higher, followed by activation of the magnetic bearing. This control means may be implemented either by the magnetic bearing control means or by a means separate therefrom. The number of revolutions of the rotor can be obtained from a rotation sensor suitably provided for.
Since when the spindle is driven at a high speed the pressure within the bearing gap of the externally pressurized gas bearing and a distribution of such pressure fluctuate, there is the possibility that the zero point of the pressure sensor may be slightly offset by the number of revolutions. If in such case the magnetic bearing is operated regularly, the predefined center of the magnetic bearing and the predefined center of the externally pressurized gas bearing may be displaced from each other within a relatively large rotational region and the direct current will flow to the electromagnet to rectify this displacement. The core loss will occur in the rotor as a result of the magnetic fluxed developed as a result of the flow of the direct current to the electromagnet, resulting in increase of the braking torque and emission of heat from the rotor to such an extent as to make it difficult to achieve a high speed rotation. Under normal conditions, it is suggested to increase the static rigidity of the magnetic bearing when rotation takes place at a rated number of revolutions or at a rate higher than the predetermined number of revolutions. In view of this, if within a predetermined rotational range, the center of rotation of the main shaft when only the externally pressurized gas bearing is activated is set to the zero point so that the magnetic bearing can be activated only when it falls within the predetermined rotational range, the rotor can be driven at a high speed. Accordingly, the spindle device utilizing this combined externally pressurized gas and magnetic bearing assembly can also be adapted for high revolution.
Yet, in the spindle device utilizing the combined externally pressurized gas and magnetic bearing assembly for rotatably supporting the main shaft having the rotor formed thereon, arrangement may be made that when the number of revolutions of the main shaft is detected as attaining a value higher than a predetermined value, the magnetic bearing control means causes a band eliminating filter to function. This system may be employed in the spindle device provided with the previously described control means at the start of the spindle device.
By inserting the band eliminating filter in a control system of the magnetic bearing which matches with the number of revolutions of the main shaft at a speed higher than the predetermined number of revolutions, a synchronizing component of the rotation of the main shaft, which forms a principal noise component of the displacement measuring means, can be eliminated to thereby limit the operation of the magnetic bearing to the low frequency. By suppressing the control at the high frequency region which is unnecessary to the magnetic bearing, the consumption of the electric power can be minimized advantageously.
Moreover, in the spindle device utilizing the combined externally pressurized gas and magnetic bearing assembly for rotatably supporting the main shaft having the rotor formed thereon, arrangement may be made that a control gain of the magnetic bearing can be lowered during a low speed rotation, but is varied to a predetermined value when the number of revolutions of the main shaft attains a value higher than a predetermined value. This setting and change of the control gain may be implemented by, for example, the previously discussed magnetic bearing control means.
Since the magnetic bearing is activated with the gain lowered in this way, any possible external disturbance to the main shaft and the rotor at the moment the magnetic bearing is activated can be suppressed advantageously. Even the setting and change of this control gain may be implemented in the spindle device of any of the previously described structures.